Hybrid Compressor System and Methods

ABSTRACT

An apparatus ( 20 ) has a centrifugal compressor ( 24 ), a positive displacement compressor ( 26 ), a first heat exchanger ( 32 ), and a second heat exchanger ( 36 ). A plurality of valves ( 70, 72, 74 ) are positioned to provide operation in at least two modes. In a first mode, refrigerant is compressed in the positive displacement compressor and the centrifugal compressor at least partially in parallel. In a second mode, refrigerant is compressed in the positive displacement compressor and the centrifugal compressor is offline. In a third mode, refrigerant is compressed in the positive displacement compressor and the centrifugal compressor at least partially in series.

CROSS-REFERENCE TO RELATED APPLICATION

Benefit is claimed of U.S. patent application Ser. No. 61/491,515, filed May 31, 2011, and entitled “Hybrid Compressor System and Methods”, the disclosure of which is incorporated by reference herein in its entirety as if set forth at length.

BACKGROUND

The disclosure relates to refrigeration. More particularly, the disclosure relates to chiller systems.

Large water-cooled chillers (e.g., 300-1500 ton capacity (1055-5275 W)) typically use a single centrifugal compressor for cost reasons. However, such compressors are subject to surge (particularly at partial load).

One way of addressing this has been to provide multiple centrifugal compressors in parallel. This allows individual compressors to be taken off line to better match compressor capacity to required load.

SUMMARY

One aspect of the disclosure involves an apparatus having a centrifugal compressor, a positive displacement compressor, a first heat exchanger, and a second heat exchanger. A plurality of valves are positioned to provide operation in at least two modes. In a first mode, refrigerant is compressed in the positive displacement compressor and the centrifugal compressor at least partially in parallel. In a second mode, refrigerant is compressed in the positive displacement compressor and the centrifugal compressor is offline. In a third mode, refrigerant is compressed in the positive displacement compressor and the centrifugal compressor at least partially in series.

In various implementations, the positive displacement compressor may be a screw compressor.

Further aspects of the disclosure involve operating in such at least two modes.

The details of one or more embodiments are set forth in the accompanying drawings and the description below. Other features, objects, and advantages will be apparent from the description and drawings, and from the claims.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a schematic view of a chiller system.

FIG. 2 is a longitudinal vertical schematic view of a condenser of the system of FIG. 1.

FIG. 3 is a longitudinal vertical schematic view of a cooler of the system of FIG. 1.

FIG. 4 is a plot of entering condenser water temperature against percent capacity.

FIG. 5 is a control flowchart for the system of FIG. 1.

Like reference numbers and designations in the various drawings indicate like elements.

DETAILED DESCRIPTION

In addition to suffering from surge problems, centrifugal compressors are typically less effective than positive displacement compressors at providing high head. This renders centrifugal compressors as poor candidates for heat reclaim operation. As is discussed below, a hybrid system features a centrifugal compressor and a positive displacement compressor. The exemplary positive displacement compressors are two- or three-rotor screw compressors powered by variable frequency drives. Alternative positive displacement compressors include reciprocating compressors and scroll compressors.

FIG. 1 shows a vapor compression system 20 having a compressor subsystem 22. The compressor subsystem 22 includes a first compressor 24 (centrifugal) and a second compressor 26 (positive displacement). In the exemplary implementation, both compressors have capacity control features which may be of any well known type (e.g., variable inlet guide vanes for the centrifugal compressor and a slide valve for a screw compressor used as the positive displacement compressor and variable speed drives for both compressors).

In a first mode of operation, the compressor subsystem drives refrigerant in a downstream direction 500 along a refrigerant flowpath 30. The flowpath 30 passes, sequentially, through a first heat exchanger 32, an expansion device 34, and a second heat exchanger 36. In the first mode, the first heat exchanger 32 is a heat rejection heat exchanger and the second heat exchanger 36 is a heat absorption heat exchanger.

An exemplary system 20 is a chiller system wherein the first heat exchanger 32 is a liquid-cooled condenser or gas cooler and the second heat exchanger 36 is the cooler. An exemplary expansion device 34 is an electronic expansion valve (EV) which may be controlled by the chiller's controller 40 (e.g., a computer or microcontroller). An alternative expansion device 34 is a float valve within the condenser 32. The exemplary first heat exchanger has at least one inlet 50 and at least one outlet 52 along the refrigerant flowpath 30. Similarly, the second heat exchanger 36 has at least one inlet 54 and at least one outlet 56 along the refrigerant flowpath. The compressor 24 has an inlet port 60 and an outlet port 62. Similarly, the second compressor 26 has an inlet port 64 and an outlet port 66.

As is discussed further below, the compressor subsystem includes one or more valves coupled to the compressors to allow switching of the compressors between two or more compression modes. The exemplary system includes three valves 70, 72, and 74.

In a first mode of operation, the compressors are operated at least partially in parallel. In an exemplary illustrated fully parallel situation, the respective suction ports and discharge ports of the compressors are at essentially identical conditions. To provide such parallel operation, the exemplary flowpath has two parallel branches 80 and 82 diverging at a junction 84 downstream of the second heat exchanger outlet 56 and re-merging at a location 86 at or upstream of the first heat exchanger inlet 50. In partially parallel situations, the separation and/or rejoinder may be at different locations.

A bypass branch or line 90 extends between the branches 80 and 82. The exemplary bypass branch 90 extends between upstream of one of the compressors to downstream of the other. In the exemplary implementation, the branch extends from a location 92 downstream of the first compressor to a location 94 upstream of the second compressor. For controlling flow in the branches 80 and 82, the exemplary valves 70 and 74 are respectively along such branches. In an exemplary implementation, the valve 70 is downstream of the first end 92 of the bypass line 90 and the valve 74 is upstream of the end 94.

In the exemplary at least partially parallel operation, the valve 72 is closed whereas the valves 74 and 70 are open. In a second mode, only the second compressor 26 is in operation. The valves 70 and 72 are closed whereas the valve 74 is open. A third mode is a series mode wherein the compressors are operated in series. In the exemplary series mode, the valves 70 and 74 are closed whereas the valve 72 is open. Refrigerant passes without diversion from the second heat exchanger outlet 56 through the first compressor, the valve 72, and the second compressor before entering the first heat exchanger inlet 50. A fourth possible mode involves having only the first compressor 24 in operation. In this mode, the valves 72 and 74 are closed and the valve 70 is open.

FIG. 1 shows further exemplary details of the condenser 32 and cooler 36. The exemplary condenser 32 includes an upper condenser tube bundle 120 and a lower subcooler tube bundle 122. FIG. 1 also shows a liquid refrigerant accumulation 124 within the condenser. The tube bundles 120 and 122 are connected to one or more sources of heat transfer fluid to withdraw heat from the refrigerant. The sub-cooler tube bundle 122 is contained within a chamber 126. One or more inlet orifices 128 are along the bottom of the chamber 126. A float valve 130 feeds the outlet 52. A pressure sensor 132 may be located in the headspace of the condenser near the inlet 50.

In an exemplary implementation, the heat transfer fluid (e.g., water) passes along a water loop 138 (FIG. 2) and is received via an inlet 140 and discharged from an outlet 142. Respective temperature sensors 144 and 146 measure inlet temperature T_(1COND) and outlet temperature T_(2COND) of the water. An exemplary flow meter 147 along the water loop 138 measures a flow rate F_(MCOND) of the water.

The cooler 36 also includes a lower tube bundle 160 and an upper tube bundle 162. FIG. 1 further shows a refrigerant accumulation 164 in the cooler. In an exemplary implementation, a heat transfer fluid (e.g., water) passes along a water loop 168 (FIG. 3) and is received via an inlet 170 and discharged from an outlet 172. Respective temperature sensors 174 and 176 measure inlet temperature T_(1COOL) and outlet temperature T_(2COOL) of the water. An exemplary flow meter 177 along the water loop 168 measures a flow rate F_(MCOOL) of the water. FIG. 1 further shows a distributor 180 in the lower portion of the cooler approximately fed by the inlet 54. A pressure sensor 182 is shown in the headspace near the outlet 56.

FIG. 4 shows a plot of the entering condenser water temperature T_(1COND) against capacity. Line 200 represents the American Refrigeration Institute (ARI) load line. In the United States, chillers are subject to ARI Standard 550. This standard identifies four reference conditions characterized by a percentage of the chiller's rated load (in tons of cooling) and an associated condenser water inlet/entering temperature. Operation is to achieve a chilled water outlet/leaving temperature of 44 F(6.67 C). The four conditions are: 100%, 85 F (29.44 C); 75%, 75 F (23.89 C), 50%, 65 F (18.33 C); and 25%, 65 F (18.33 C also). These conditions (or similar conditions along a curve of connecting them) may provide relevant conditions for measuring efficiency. In API testing, the water flow rate through the cooler is 2.4 gallons per minute per ton of cooling (gpm/ton) (0.043 liters per second per kilowatt (l/s/kW)) and condenser water flow rate is 3 gpm/ton (0.054 l/s/kW).

With typical heat exchangers, the water temperature rise across the condenser is approximately 8F (4.4 C) times the percentage load or 8 F at 100% load, 6F (3.3 C) at 75% load, 4F (2.2 C) at 50% load and 2F (1.1 C) at 25% load. The cooler saturation temperature is 1F (0.6 C) or 2F (1.1 C) below the leaving chilled water temperature (e.g., 43 F in the ARI test). Similarly, the condenser saturation temperature is 1 F or 2 F above the sum of entering condenser water temperature and water temperature rise. For entering condenser water temperature of 85 F, temperature rise is 8 F, hence leaving condenser water temperature is 93 F and condenser saturation temperature is 93+2=95 F.

Line 202 represents a constant temperature of 85 F (29.44 C). In tropical regions, the ambient temperature changes very little from day to night. In such regions, the condenser water temperature remains constant. It's an industry standard, to consider the entering condenser water temperature constant at 85 F between 25% and 100% load. Table I shows lift for the ARI conditions and corresponding tropical conditions. Centigrade temperatures are conversions from the listed Fahrenheit values and thus do not add and present false precision. Other SI parentheticals herein similarly represent conversions from the original US or English values.

TABLE I ARI entering Constant entering ARI condenser water condenser water Constant entering Load temperature temperature condenser water lift (%) (F.(C.)) ARI lift (F.(C.)) (F.(C.)) (F.(C.)) 100 85 (29.44) 95 − 43 = 52 85 (29.44) 95 − 43 = 52  (35 − 6.11 = 28.89)  (35 − 6.11 = 28.89) 75 75 (23.89) 82.5 − 43 = 39.5 85 (29.44) 92.5 − 43 = 49.5 (28.06 − 6.11 = 21.94)  (33.61 − 6.11 = 27.5)   50 65 (18.33) 70 − 43 = 27 85 (29.44) 90 − 43 = 47 (21.11 − 6.11 = 15)   (32.22 − 6.11 = 26.11)  25 65 (18.33) 67.5 − 43 = 24.5 85 (29.44) 87.5 − 43 = 44.5 (19-72 − 6.11 = 13.61)  (30.83 − 6.11 = 24.72) 

In a first exemplary implementation, the at least partially parallel first mode is utilized at high loads and the second mode (positive displacement compressor-only) is used at low loads. For example, the second mode may be used from essentially zero load to an intermediate load value. Between the intermediate load value and the maximum load, the at least partially parallel mode is used. It is noted that the intermediate load value may, however, be subject to appropriate hysteresis control to avoid excessive cycling when operating near changeover conditions. For example, from zero to changeover, the second compressor may be operated at increasing speed and/or power. At changeover, the first compressor may be brought online at full or near full capacity and the second compressor reset to zero or other low capacity value. Thereafter, with increasing load, the speed and/or power of the second compressor may be increased. Thus, in the at least partially parallel mode system load/capacity variations are principally accommodated by the positive displacement compressor. For example, at a minimum, the positive displacement compressor may address more of the variation than the centrifugal compressor does. More narrowly, the positive displacement compressor may address at least 75% or at least 90% of the load variation. The load variation may represent at least an exemplary 30% of a peak load of the system, more narrowly, at least 40%.

In the first, simple, exemplary implementation, the rated capacities of the two compressors are essentially the same (e.g., the same or appropriately differently sized to address any hysteresis issues). The changeover point is, therefore, at essentially half load.

More broadly, the changeover point may be between 45% and 55% or 40% and 60% of the total rated system load.

By using the centrifugal compressor only at or near its own rated load (or, more broadly, not at a low load) issues of surge may largely be avoided.

For example, an exemplary rated maximum capacity of the positive displacement compressor is 50-200% of the rated maximum capacity of the centrifugal compressor, more narrowly, 100 to 150%.

In one variation, the fourth (series) mode may be added and used at high condenser water temperatures such as a water heating or a heat reclaim mode.

In a further variation, the centrifugal compressor may be used alone when very low lift is needed (e.g., less than 25 F (13.9 C)).

In a second, more complex, exemplary implementation, a control process 300 (FIG. 5) starts by measuring or otherwise determining 302 the saturation temperatures of the condenser (T_(COND)) and the cooler (T_(COOL). T) _(COND) and T_(COOL) may respectively be determined by measuring the pressures via the pressure sensors 132 and 182 and then calculating the saturation temperatures (either via a lookup table or programmed function). The lift is calculated 304 as T_(COND) minus T_(COOL). If the lift is greater than a given threshold (e.g., 50 F (28 C)) the system may be operated 306 in the fourth (series) mode. In the series mode, the capacity of the centrifugal compressor is controlled by compressor speed and by inlet guide vane orientation. If the volume of discharge gas from the centrifugal compressor is higher than the capacity of the screw compressor, then the pressure between the centrifugal compressor and the screw compressor will rise resulting in surge of the centrifugal compressor. At this point, the centrifugal compressor speed is incrementally increased and its guide vanes are incrementally closed until the centrifugal compressor comes out of surge. A similar logic is applied for the screw compressor (i.e., first speed followed by slide valve). Reducing the speed always results in reduced power consumption or increased efficiency.

For lesser lift, a non-series operation may be performed. Measurements 308 are made of the flow rate F_(MCOOL) and the temperatures T_(1COOL) and T_(2COOL). Capacity may also be calculated 310.

At low capacity (e.g. less than a first value such as 50% of a maximum) operation is then refined based upon the head. For low head, the compressors may be run 320 in the first mode at equal loads. This may involve controlling capacity via the speed when variable speed drive is present and by the centrifugal compressor inlet guide vanes and the screw compressor slide valve for fixed speed case. Head is proportional to the temperature lift. In the example, low head corresponds to temperature lift less than 35 F (19 C) and high head between 35 F and 50 F (19 C and 28 C).

For high head, the system is operated 322 in the second mode (screw compressor-only). Capacity is controlled via speed when variable speed drive is present and by slide valve for fixed speed case.

At intermediate capacity (e.g., 50-75%), operation may also be in the first mode. Balance between the compressors may be refined based upon the head. For low head, the system is run in the parallel mode with the screw compressor operating at a fixed capacity and the centrifugal compressor operating at variable capacity to provide the required overall capacity. For example, with centrifugal and screw compressors of maximum equal capacity, the screw compressor may be operated at 50% of its maximum capacity and the centrifugal compressor being operated at between 50 and 100% of its maximum capacity (thereby combining to provide the exemplary 50-75% of maximum system capacity operation). Such an operation is chosen so as to avoid surge of the centrifugal compressor.

At high head, the system may be run 332 in the first mode with the centrifugal compressor at essentially fixed capacity and the screw compressor providing capacity control. The exemplary set points of the constant capacity compressor may differ relative to the condition 330. The centrifugal compressor may be run at a relatively high capacity. In the foregoing example, this may be at 80% at its maximum capacity thereby providing 40% of total system capacity. The screw compressor may be run between 20 and 70% of its maximum capacity (thereby providing 10-35% of maximum system capacity and combining with the centrifugal compressor to provide the 50-75% of maximum system capacity). Such operating condition may be selected because the centrifugal compressor is susceptible to surge at low loads and high head.

At high capacity, (e.g., 75-100%), operation may also be in the first mode, with compressors running 340 at equal loads. Thus, depending upon the necessary capacity, each may be run at between 75 and 100% of its own maximum capacity to satisfy the required capacity.

One or more embodiments have been described. Nevertheless, it will be understood that various modifications may be made. For example, in reengineering an existing system, details of the existing system or its intended use may influence details of any particular implementation. Accordingly, other embodiments are within the scope of the following claims. 

1. An apparatus (20) comprising: a centrifugal compressor (24); a positive displacement compressor (26); a first heat exchanger (32); a second heat exchanger (36); and a plurality of valves (70, 72, 74) positioned to provide at least two of: operation in a first mode wherein: refrigerant is compressed in the positive displacement compressor and the centrifugal compressor at least partially in parallel; operation in a second mode wherein: refrigerant is compressed in the positive displacement compressor; and the centrifugal compressor is offline; and operation in a third mode wherein: refrigerant is compressed in the positive displacement compressor and the centrifugal compressor at least partially in series.
 2. The apparatus of claim 1 wherein: the positive displacement compressor is a screw compressor.
 3. The apparatus of claim 1 wherein: the centrifugal compressor has a rated maximum capacity; and the positive displacement compressor has a rated maximum capacity of 100-150% of the rated maximum capacity of the centrifugal compressor.
 4. The apparatus of claim 1 wherein: the plurality of valves are positioned to provide all three of said modes.
 5. The apparatus of claim 1 wherein: the plurality of valves are positioned to provide at least the first mode and second mode.
 6. The apparatus of claim 1 being a chiller wherein: the first heat exchanger is part of a condenser unit; the second heat exchanger is part of a cooler; and an expansion device (34) is positioned between the condenser and the cooler.
 7. The apparatus of claim 1 wherein the plurality of valves comprises: a first valve (70) between the centrifugal compressor and the first heat exchanger; a second valve (72) along a bypass (90) extending from a location (92) between the centrifugal compressor and the first valve to a location (94) between the second valve and the screw compressor; and a third valve (74) between the second heat exchanger and the positive displacement compressor.
 8. The apparatus of claim 1 further comprising: a controller (40) programmed to automatically switch between said at least two modes.
 9. The apparatus of claim 8 wherein: the controller is programmed to automatically switch between said first mode and said second mode.
 10. The apparatus of claim 9 wherein: said controller is programmed to switch from said first mode to said second mode responsive to a decrease in load and from said second mode to said first mode responsive to an increase in load.
 11. The apparatus of claim 10 wherein: the controller is programmed to switch to said third mode responsive to calculating a high requirement for lift.
 12. An apparatus comprising: a centrifugal compressor (24); a positive displacement compressor (26); a first heat exchanger (32); a second heat exchanger (36); and means (70, 72, 74) for providing at least two of: operation in a first mode wherein: refrigerant is compressed in the positive displacement compressor and the centrifugal compressor at least partially in parallel; operation in a second mode wherein: refrigerant is compressed in the positive displacement compressor; and the centrifugal compressor is offline; and operation in a third mode wherein: refrigerant is compressed in the positive displacement compressor and the centrifugal compressor at least partially in series.
 13. A method for operating a vapor compression system (20) having a centrifugal compressor (24) and a positive displacement compressor (26), the method comprising at least two of: operating in a first mode wherein: refrigerant is compressed in the positive displacement compressor and the centrifugal compressor at least partially in parallel; operating in a second mode wherein: refrigerant is compressed in the positive displacement compressor; and the centrifugal compressor is offline; and operating in a third mode wherein: refrigerant is compressed in the positive displacement compressor and the centrifugal compressor at least partially in series.
 14. The method of claim 13 wherein: the system is operated in the second mode from a minimum load condition to an intermediate load condition; and the system is operated in said first mode at loads above said intermediate load condition.
 15. The method of claim 14 wherein: in the first mode, load variation is principally accommodated by the positive displacement compressor.
 16. The method of claim 13 wherein: operation is based on a combination of required capacity and required lift and wherein some-to-all of: in a high required lift, the system is operated (306) in the third mode; and at low required lift, some-to-all of: at low capacity: the system is run (320) in the first mode; and at high head the system is run (322) in the third mode; at intermediate capacity: at low head, the system is run (330) in the third mode with essentially constant positive displacement compressor capacity and the centrifugal compressor at variable capacity; and at high head, the system is run (322) in the first mode with the centrifugal compressor at essentially constant capacity and the positive displacement compressor at variable capacity; and at high capacity, the system is run (340) in the first mode with both compressors providing variable capacity.
 17. The method of claim 13 wherein: the positive displacement compressor and the centrifugal compressor are both electrically powered by variable speed drives.
 18. The apparatus of claim 12 wherein: the positive displacement compressor and the centrifugal compressor are both electrically powered by variable speed drives.
 19. The apparatus of claim 1 wherein: the positive displacement compressor and the centrifugal compressor are both electrically powered by variable speed drives. 